Hydrostatic power transmission system

ABSTRACT

An improved hydrostatic power transmission is disclosed including first, second and third fluid displacement units, of which the first and second fluid displacement units are constructed as variable-displacement and constant-displacement units, respectively, while the third fluid displacement unit may be constructed either as a variable-displacement or as a constant-displacement unit as the case may be. The transmission system features split-power characteristics in which the second fluid displacement unit acts not only as a hydrostatic unit but as a mechanical reaction unit whereby the input power is split into two input and/or output components. The transmission efficiency is consequently increased to a significant extent in spite of the simple and small-sized construction of the transmission system. A typical practical application of the transmission system is motor vehicles including heavy-duty industrial vehicles in which powerful braking actions are required when they are to be stopped.

United States Patent 1 Mori [ June 26, 1973 HYDROSTATIC POWERTRANSMISSION SYSTEM Primary ExaminerEdgar W. Geoghegan Attorney-JohnLezdey [57] ABSTRACT An improved hydrostatic power transmission isdisclosed including first, second and third fluid displacement units, ofwhich the first and second fluid displacement units are constructed asvariable-displacement and constant-displacement units, respectively,while the third fluid displacement unit may be constructed either as avariable-displacement or as a constantdisplacement unit as the case maybe. The transmission system features split-power characteristics inwhich the second fluid displacement unit acts not only as a hydro staticunit but as a mechanical reaction unit whereby the input power is splitinto two input and/or output components. The transmission efficiency isconsequently increased to a significant extent in spite of the simpleand small-sized construction of the transmission system. A typicalpractical application of the transmission system is motor vehiclesincluding heavy-duty industrial vehicles in which powerful brakingactions are required when they are to be stopped.

PAIENIEDJUNZG I975 3.740.953

saw u or 4 INVENTOR YOICH/ OR! B 1 HYDROSTATIC POWER TRANSMISSION SYSTEMThis invention is concerned generally with power transmission systemsand has its particular reference to a hydrostatic power transmissionsystem providing steplessly variable speed ratio characteristics. Thepower transmission system herein dealt with is specifically adapted tobe used in an automotive vehicle driveline for delivering power from avehicle power plant such as an internal combustion engine or a gasturbine engine to the vehicle traction wheels. Such power transmissionsystem uses positive-displacement pistontype hydrostatic units having aclosed hydrostatic circuit to provide power delivery paths from adriving member to a driven member.

Hydrostatic power transmissions generally use a variable-displacementpump unit acting as a source of fluid power and a constant-displacementhydraulic motor unit usually driven by the pump unit. Both the pump andmotor units are of the piston type using, for instance, ball pistons andare respectively connected to suitable driving and driven members whichmay be input and output shafts of a power train of automotive vehicles.Power from the driving member or input shaft is delivered to the drivenmember or output shaft by the aid of the fluid as a result of the pumpdelivery and the fluid pressure.

The positive-displacement piston-type hydrostatic power transmissionsare adapted to provide ease and simplicity of operation because of theirability to steplessly deliver outputs of various speeds from a source ofpower having a given revolution speed and because of the fact that thestepless change of the revolution speed and even reversed motions of thehydraulic motor units can be effected without use of clutches and gearedreduction mechanisms. Less shooks and vibra tions are invited than inthe purely mechanical power transmissions and fluctuations in the torquetransmitted are largely subdued by the working fluid itself, thusreducing to a minimum the time and labour for the periodical inspectionand maintenance servicing of the transmission as a whole.

Hydrostatic power transmissions of modified types are known in which thefluid power acting upon the pump and motor units is split into two inputor output components in the hydrostatic circuit. Where the fluid poweris to be split into input components, recirculation of the power isbrought about during deceleration condition in which the vehicle drivesthe power plant, thereby causing reduction in the transmission effciency. This problem is avoided in the hydrostatic power transmission inwhich the fluid power is split into output components but, in the powertransmission of this particular type, the range of the reduction ratiosavailable is limited due to restrictions resulting from I the veryconstruction of the transmission. Since, moreover, the pressurized fluidin the hydrostatic circuit should be partially drained off when thetransmission is about to establish a neutral condition interruptingdelivery of power to the driven member, braking actions attained in thetransmission alone are insufficient to bring the vehicle to a full stop.The hydrostatic power transmission of the split-output type is, thus,not applicable to heavy-duty industrial vehicles such as bulldozers,tractors and forklift trucks.

It is, therefore, an object of this invention to provide an improvedpositive-displacement hydrostatic power transmission system of thesplit-power type which is cleared of these drawbacks inherent in theprior art hydrostatic power transmissions of the split-input orsplitoutput type.

It is, thus, another object of this invention to provide an improvedvariable-speed hydrostatic power transmission system in which the poweracting upon the hydrostatic units forming part of the transmissionsystem is split into mechanical and hydraulic components so as toprovide increased speed and torque transmission efficiencies throughoutvarious modes of operation in cluding the deceleration and neutralconditions.

It is still another object of the invention to provide an improvedvariable-speed power transmission system in which braking actions ofsufficient magnitudes can be applied on the driven member under theneutral condition of the transmission system. The hydrostatic powertransmission. having such feature is successfully applicable toheavy-duty industrial vehicles.

It is still another object of the invention to provide an improvedhydrostatic power transmission system of the 'split-powercharacteristics in which the speed and torque ratios between the drivingand driven members are varied stoplessly in either direction from zeroto infinity at satisfactorily high efficiencies under the various modesof operation.

It is still another object of the invention to provide an improvedpositive-displacement variable-speed hydrostatic power transmissionsystem in which the power transmission system in which the powersupplied thereto is split into input and/or output components withoutuse of mechanical clutches and geared reduction mechanisms.

it is still another object of the invention to provide an improvedpositive-displacement variable-speed hydrostatic power transmissionsystem which is simple in construction and operation and which iscapable of dealing with great power in spite of its relativelysmallsized construction.

It is still another object of the invention to provide an improvedhydrostatic power transmission system which can be snugly and compactlyinstalled on automotive and industrial vehicles.

These and other objects of this invention are advantageouslyaccomplished in a hydrostatic power transmission which includes first,second and third fluid displacement units cooperating with each other.Each of these fluid displacement units comprises a rotatable cylinderblock coaxial with the driving or driven member and a cam ring having aninner cam surface and positioned around the cylinder block. The cylinderblocks of the first and second fluid displacement units are rotatablewith a driving member and the cam ring of the second fluid displacementunit and the cylinder block of the third fluid displacement unit arerotatable with or in driving engagement with a driven member. Each ofthe cylinder blocks has a plurality of cylinders which are directedtoward an axis of rotation of the cylinder block and a plurality ofpiston elements which are respectively received within these cylindersin a manner to be movable toward and away from the axis of rotation ofthe cylinder block. The piston elements are in frictional engagementwith the inner cam surface of the associated cam ring. The cam ring ofthe first fluid displacement unit is movable over the associatedcylinder block in a plane transverse to the axis of rotation of thecylinder block for providing controlled degrees of eccentricity betweenthe inner cam surface of the particular cam ring and the above mentionedaxis. Associated with this cam ring of the first fluid displacement unitis actuating means which is adapted to displace the cam ring at a rightangle to the axis of rotation of the associated cylinder block forvarying the displacement of the first fluid displacement unit inaccordance with desired operating conditions. The cam ring of the secondfluid displacement unit, on the other hand, is held stationary so thatthe displacement of the second fluid displacement unit is kept constant.Fluid pressure distribution passage means provides controlled fluidcommunication between the cylinders of the first, second and third fluiddisplacement units for radially outwardly biasing the piston elementsreceived therewithin. The first fluid displacement unit is operable todisplace the fluid in either direction from zero to maximum and themaximum displacement per turn of the cylinder block thereof is greaterthan the displacement of the second displacement unit. The third fluiddisplacement unit is operable to displace the fluid in a fixed directionwhich usually corresponds to the direction of rotation of the drivingmember.

Under an underdrive condition in which the driven member is to be drivenat a speed relatively lower than and in the same direction as thedriving member, the adjustable cam ring of the first fluid displacementunit is moved relative to the associated cylinder block to a position inwhich the displacement thereof is smaller than the displacement of thesecond fluid displacement unit. In this condition, the displacement ofthe second fluid displacement unit is split into two components, one ofwhich acts upon the third fluid displacement unit to increase the torqueon the driven member and the other of which acts upon the first fluiddisplacement unit so as to increase the torque on the input shaft. Thetransmission system thus operates in such a manner as to have its fluidpower split into input and output components under the underdrivecondition.

The first fluid displacement unit is constructed as avariable-displacement unit while the second fluid displacement unit isconstructed as a constantdisplacement unit. The third fluid displacementunit, on the other hand, may be either a variabledisplaccment unit or aconstant-displacement unit as the case may be. Where the third fluiddisplacement unit is constructed as the variable-displacement unit withits cam ring rockable with respect to the associated cylinder block, theunit is adjusted to provide a maximum displacement during the neutral,underdrive and reverse drive conditions, in which the displacement ofthe first fluid displacement alone is varied from minus to plus valuesthrough zero whereby the power is split in different manners. Underdeceleration condition with the driven member driven at considerablyhighspeeds, the cam ring of the first fluid displacement unit is movedto be in concentric alignment with the axis of rotation of theassociated cylinder block so that the delivery of the fluid therefrom isinterrupted. In this condition, only the displacement of the third fluiddisplacement unit is varied or decreased so that the fluid power issplit into two output components. When, furthermore, the driven memberis to be coupled with and driven at the same speed and in the samedirection as the driving member to establish a direct drive condition,the displacements of the first and third fluid displacement units areinterrupted concurrently. The secorid fluid displacement unit isconsequently brought into a so-called hydraulic lock condition so thatthe transmission system operates as an integral unit. To establish anoverdrive condition, the displacement of the third fluid displacementunit is interrupted and at the same time the direction of thedisplacement of the first fluid displacement unit is reversed, the fluidpower thus being split into input components.

Where, on the other hand, the third fluid displacement unit isconstructed as the constant-displacement unit with its cam ring fixedrelative to the associated cylinder block, the transmission system nowacts as a so-called hyperbolic system in which the over-all speedreduction ratio that can be achieved is related to the ratio with ahyperbolic function containing only one variable which is thedisplacement of the first fluid displacement unit. In this instance, thedriven member stops when the displacement of the first fluiddisplacement unit is equalized with that of the second fluiddisplacement unit. As the displacement of the first displacement unitbecomes smaller than the displacement of the second fluid displacementunit, then the driven member is driven in the same direction as thedriving member. As the displacement of the first fluid displacement unitexceeds the displacement of the second fluid displacement unit, thedriven member is driven in the reverse direction. When the direction ofthe displacement of the first fluid displacement unit is reversed andthe absolute value of the displacement is equalized with thedisplacement of the third fluid displacement unit, then the drivenmember is coupled with the driving member so that the two membersoperate at the same speed. When, furthermore, the absolute value of thedisplacement of the first fluid displacement unit is greater than theabsolute value of the displacement of the third fluid displacement unit,then an overdrive condition is established in the transmission system.

The fluid pressure distribution passage means may be provided by axialbores formed in the driving and driven members which are in line witheach other. The bores in the driving and driven members communicate witheach other at their respective ends and with the cylinders of the first,second and third fluid displacement units through ports and groovesformed in the driving and driven members and in sleeves mounted on outerperipheral surfaces of the members. Such fluid pressure distributionpassage means is occupied with a fluid under pressure so that the ballpiston elements in the fluid displacement units are constantly biasedradially outwardly.

The aligned bores in the driving and driven members are closed at theiropposed ends. Where, however, the third fluid displacement unit isconstructed as the variable-displacement unit with its cam ringpositionadjustable with respect to the associated cylinder, the bores inthe driving and driven members may be provided with extensions at theiropposed ends so as to accommodate therein a pair of valve means whichare respectively associated with the first and third fluid displacementunits having the variable-displacement char acteristics. When the camrings of the first and third fluid displacement units are adjusted sothat the displacements thereof are interrupted, these valves are causedto block entry of the highly pressurized fluid into the cylinders ofthese fluid displacement units. Thus, even though the cylinders of thefirst and third fluid displacement units are open to the fluid pressuredistribution passage means due to variation in the angular positions ofthe cylinder blocks relative to the passage means, pulsating fluidpressures are not imparted to the first and third fluid displacementunits. Production of noises otherwise produced under the direct drivecondition of the transmission system is avoided in this manner and, ontop of this, leakage of the fluid from the first and third fluiddisplacement units is minimized so as to save the fluid distributionefficiency from being degraded.

The modes of operation of the transmission system thus constructed canbe varied steplessly simply by varying the angular position of thevariabledisplacement unit or units of the transmission system by the useof the actuating means which is adapted to be responsive to the desiredoperating conditions. Only a limited amount of reaction is imparted tosuch actuating means during operation because the cam ring of thevariable-displacement unit is arranged to be readily rockable around theassociated cylinder block.

The features and advantages of the hydrostatic power transmission systemimplementing this invention will be more clearly understood from thefollowing description and from the accompanying drawings, in which:

. FIG. 1 is a lengthwise sectional view showing a preferred embodimentof the hydrostatic power transmission system according to thisinvention;

FIG. 2 is a diagrammatic view showing, on a translatory basis,fundamental modes of operation of the hy' drostatic power transmissionillustrated in FIG. 1;

FIG. 3 is a view similar to FIG. 1 but now shows another preferredembodiment of the power transmission system of this invention; and

FIG. 4 is a view graphically demonstrating performance characteristicsof the power transmission system shown in FIG. 3.

While the embodiments of the system according to this invention will bedescribed as being of the character using ball pistons, such is solelyfor illustrative purposes and it should be borne in mind that the gistof the invention is applicable to the hydrostatic transmissions usingswash plates and axial pistons or vaned pumps.

Reference is now made to the drawings, first to FIG. 1. The hydrostaticpower transmission system as shown intervenes between an input shaft asa driving member and an output shaft 11 as a driven member so that thepower is delivered from the input shaft 10 to the output shaft 11 by thetransmission system. Where the transmission system is incorporated in amotor vehicle, the input shaft 10 is connected to and driven by acrankshaft of a vehicle power plant such as an internal combustionengine or a gas turbine engine and the output shaft 11 is connected tothe vehicle traction wheels via a suitable driveline, though not shown.The transmission system as a whole is encased in a transmission housingwhich is generally indicated by reference numeral 12 in FIG. 1.

The hydrostatic power transmission system is essentially made up offirst, second and third fluid displacement units 13, 14 and 15,respectively. The first and third fluid displacement units 13 and 15,respectively, are constructed as variable-displacement units while thesecond fluid displacement unit 14 constructed as a constant-displacementunit.

The first fluid displacement unit 13 includes a cylinder block 16 whichis rotatable with the input shaft 10 through a key 17 interposedtherebetween. If desired,

the cylinder block 16 may be splined to the input shaft 10 without useof the key 17, though not shown as such. The cylinder block 16 issupported on the transmission housing 12 through an annular bearing 18.This cylinder block 16 has a generally circular cross section and isformed with a suitable number of substantially equidistantly spacedpiston cylinders which are shown to be constituted by cylindricalopenings or chambers of which only two are seen in FIG. 1 as indicatedby reference numerals 19a and 19b. These cylindrical chambers are alldirected toward an axis of rotation of the cylinder block 16 and thecylindrical chambers 19a and 19b as shown are assumed to besubstantially diametrically opposed to each other. Ball piston elements20a, 20b, are received within the cylindrical chambers 19a, 19b,respectively, in a manner to be slidable toward and away from the axisof rotation of the cylinder block 16. The cylinder block 16 issurrounded by an adjustable cam ring 21 having a circular inner camsurface. The individual ball piston elements 20a, 20b, are in frictionalengagement with this inner cam surface of the cam ring 21. This innercam surface may be formed with an annular groove 22 thereby to add tothe area of frictional contact between the ball piston ele ments and thecam ring. The cam ring 21 is pivotally supported by the transmissionhousing 12 through a pin 23 so as to be rockable around the peripheralwall of the cylinder block 16 in a plane transverse to the axis ofrotation of the cylinder block. The cam ring 21 thus beingposition-adjustable with respect to the associated cylinder block 16,controlled degrees of eccentricity are defined between the cam ring andthe axis of rotation of the cylinder block depending upon the angularpositions of the cam ring. This angular position of the cam ring isvaried by suitable actuating means, not shown, which is operativelyconnected to the cam ring through a pin 24. This actuating means iseither manually or automatically operated in accordance with the desiredoperating conditions of the motor vehicle, for instance. Where theactuating means is to be automatically operated, the means may include asuitable fluid operated servo mechanism responsive to the operatingconditions of the motor vehicle and connected to the pin 24 of the camring 21 through a piston rod 25 as shown.

The second fluid displacement unit 14 includes a cylinder block 26 whichis rotatable together with the cylinder block 16 of the first fluiddisplacement unit 13 and on the output shaft 11. This cylinder block 26is thus illustrated as integral with the cylinder block 16 of the unit13. Similarly to the cylinder block 16, the cylinder block 26 has acircular cross section and is formed with a plurality of radiallyextending spaced cylindrical chambers 27a, 27b, of which the chambers27a and 27b as seen in FIG. 1 are assumed to be diametrically opposed toeach other. Ball piston elements 28a, 28b, are received within thecylindrical chambers 27a, 27b, respectively, in a manner to be slidabletoward and away from an axis of rotation of the cylinder block 26. A camring 29 is eccentrically positioned around the cylinder block 26 and hasa circular inner cam surface with which the ball piston elements 28a,28b, are constantly in frictional engagement. The degree of theeccentricity between the inner cam surface of the cam ring 29 and theaxis of rotation of the cylinder block 26 is kept constant so that theball piston elements 28a, 28b, reciprocate within the cylindricalchambers 27a, 27b, at all times in a regular manner. The cam ring 29 mayhave formed in its inner cam surface an annular groove 30 so as to addto the area of friction between the ball piston elements 28a, 28b, andthe cam ring 29. The second fluid displacement unit 14 is in this mannerconstructed to act as a constant displacement unit.

The third fluid displacement unit of the transmission system illustratedin FIG. 1 is essentially similar in construction to the first fluiddisplacement unit 13, constructed as the variable displacement unit.Thus, the third fluid displacement unit 15 includes a cylinder block 31which is rotatable with the output shaft 11 through a key 32 interposedtherebetween. The cylinder block 31 is integral with the cam ring 29 ofthe second fluid displacement unit 14 so that the cylinder block 31 andthe cam ring 29 are rotatable together. The cylinder block 31 issupported by the transmission housing 12 through an annular bearing 33and the rotation thereof is isolated from the rotation of the cylinderblock 26 of the second fluid displacement unit 14 by means of an annularbearing 34 which is interposed between these cylinder blocks 26 and 31.A plurality of spaced cylindrical chambers 35a, 35b, are formed radiallyin the cylinder block 31, of which the two cylindrical chambers 35a and35b are herein assumed to be diametrically opposed to each other. Thecylindrical chambers a, 35b, receive therewithin ball piston elements36a, 36b, respectively, which are slidable toward and away from an axisof rotation of the cylinder block 31. An adjustable cam ring 37 having acircular inner cam surface is positioned around the cylinder block. Thecam surface of this cam ring 37 may preferably be formed with an annulargroove 38 for the reason previously discussed. The cam ring 37 ispivotally supported by the transmission housing 12 through a pin 39 soas to be rockable with respect to the associated cylinder block 31 in aplane which is substantially at a right angle to the axis of rotation ofthe cylinder block, whereby controlled degrees of eccentricity can beestablished between the inner cam surface of the cam ring 37 and theaxis of rotation of the cylinder block 31. The cam ring 37 isposition-adjusted with respect to the cylinder block 31 by suitableactuating means, not shown, to which the cam ring may be operativelyconnected through one pin 40. This actuating means may be constructedessentially similarly to its counterpart in the first fluid displacementunit 13 and, where such means is to include a fluid operated servomechanism as previously mentioned, the servo mechanism may be connectedto the pin 40 for the cam ring 19 through a piston rod 41 asillustrated.

Fluid communication is provided between the cylindrical chambers of thefirst, second and third fluid displacement units 13, 14 and 15,respectively, through pressure distribution passage means, wherebytorques are transmitted therebetween under various modes of operation ofthe transmission system. For this purpose, the cylinder blocks 16, 26and 31 have formed in their inner peripheral walls radial ports 46a,46b, 47a, 47b, and 48a, 48b, which emerge from the cylindrical chambers19a, 19b, 27a, 27b, and 35a, 35b, respectively, and open toward theinput and output shafts l0 and 11. A sleeve 49 surrounds that portion ofthe input shaft 10 which is approximately coextensive with the cylinderblock 16 of the first fluid displacement unit 13. This sleeve 49 is heldstationary relative to the transmission housing 12. A pair ofsemi-circumferential grooves 501: and 50b are formed in an outerperipheral wall of the sleeve 49. These circumferential ports 50a and50b are diametrically opposed to each other and are periodically incommunication with the radial ports 46a, 46b, as the cylinder block 16is rotated with the input shaft 10. The semi-circumferential grooves 50aand 50b communicate with elongated grooves 51a and 51b, respectively,which are formed longitudinally in the outer peripheral wall of thesleeve 49. This sleeve 49 is further formed with radial ports 52a and52b which lead from the elongated grooves 510 and 51b, respectively.

The input and output shafts 10 and 11, respectively, are joined to eachother through spigot connection as at 43 in such a manner that theoutput shaft 11 is free from the rotation of the input shaft 10. Theinput and output shafts 10 and 11 have formed therein axial bores 54 and55, respectively, which meet each other at t;,. Thus, the axial bore 54is opened at the terminal end of the input shaft 10 while the axial bore55 is opened at the starting end of the output shaft 11. These axialbores 54 and 55 as a whole are at least partly coextensive with anassembly of the first, second and third fluid displacement units 13, 14and 15, respectively, and are closed at their opposed ends in a suitablemanner which will be described later.

The input shaft 10 has formed in its outer peripheral wall a pair ofspaced annular grooves 56a and 56b which are aligned with the radialports 52a and 5212, respectively, in the sleeve 49. The annular groove56a is in communication with the axial bore 54 in the input shaft 10through a radial passage 57 which is formed in the input shaft asillustrated. The annular groove 56b, on the other hand, communicateswith an elongated groove 58 which is formed longitudinally in the outerperipheral wall of the input shaft 10 and which terminates at the end ofthe input shaft close to the output shaft 11. Thus, as the input shaft10 and the cylinder block 16 mounted thereon are rotated with respect tothe stationary sleeve 49, fluid communication is periodicallyestablished between the cylindrical chambers 19a and 19b in the cylinderblock 16 and the bore 54 and elongated groove 58 in the input shaft 10.More specifically, the communication between the chamber 19a and theaxial bore 54 is provided through the radial port 46a in the cylinderblock, the semi-circumferential groove 50a, elongated groove 51a andradial port 52a in the sleeve 49, and the annular groove 56a and passage57 in the input shaft 10 The communication between the chamber 19b andthe elongated groove 58, on the other hand, is provided through theradial port 46b in the cylinder block 16, the semi-circumferentialgroove 50b, elongated groove 51b and radial port 52b in the sleeve 49,and the annular groove 56b in the input shaft 10.

The output shaft 11 has formed in its outer peripheral wall close to theinput shaft 10 a pair of substantially diametrically opposedsemi-circumferential grooves 59a and 59b which are aligned with theradial ports 47a, 47b, in the cylinder block 26 of the second fluiddisplacement unit 14. The semi-circumferential groove 59a, which isshown as instantly in alingment with the radial port 47a, communicateswith the axial bore 55 in the output shaft through a radial passage 60formed in the output shaft. The semi-circumferential groove 59b, whichis shown as instantly in alignment with the radial port 47b, emergesfrom the elongated groove 58 in the input shaft 10 through an annulargroove 61 which is formed between joined end walls of the input andoutput shafts 10 and 1 1, respectively. The semi-circumferential groove59b leads to an elongated groove 62 which is formed longitudinally inthe outer peripheral wall of the output shaft 11. The axial bore 55 andthe elongated groove 62 formed in the output shaft 11 are in this mannerperiodically permitted to communicate with the cylindrical chambers 27a, 27b, as the cylinder block 26 rotates relative to the output shaft 11.

The output shaft 11 has its portion approximately coextensive with thecylinder block 31 of the third fluid displacement unit surrounded by asleeve 63 which is stationary relative to the transmission housing 12.The sleeve 63 has formed in its outer peripheral wall a pair ofdiametrically opposed semi-circumferential grooves 64a and 64b which areperiodically in communication with the radial ports 48a, 48b, in thecylinder block 31 as this cylinder block rotates with respect to thesleeve. The semi-circumferential grooves 640 and 64b merge intoelongated grooves 65a and 65b, respectively, which are formedlongitudinally in the outer peripheral wall of the sleeve 63. The sleeve63 is further formed with radial ports 66a and 66b which emerge from theelongated grooves 65a and 65b, respectively.

Close to the cylinder block 31 of the third fluid displacement unit 15have formed in the outer peripheral wall of the output shaft 11 a pairof spaced annular grooves 67a and 67b which are aligned with the radialports 66a and 66b, respectively, in the sleeve 63. The annular groove67a, which is shown as instantly in corrimunication with the radial port66a, communicates with the axial bore 55 in the output shaft 11 througha radial passage 68 formed in the output shaft. The other annular groove67b, which is shown as instantly in communication with the radial port66b, communicates with the semi-circumferential groove 59b for thesecond fluid displacement unit 14 through an elongated groove 62 formedlongitudinally in the outer peripheral wall of the output shaft 11. Thiselongated groove 62 is at all times in communication with the elongtedgroove 58 in the input shaft 10 through provision of the annular groove61 between these elongated grooves. Thus, as the output shaft 11 and thecylinder block 31 mounted thereon are rotated together with respect ofthe stationary sleeve 63, fluid communication is periodicallyestablished between the cylindrical chambers a and 35b and the axialbore and the elongated groove 62, respectively, in the output shaft 11.The communication between the chamber 35a and the axial bore 55 isprovidedthrough the radial port 48a in the cylinder block 31, thesemi-circumferential groove 64a, elongated groove 65a and radial port66a in the sleeve 63, and the annular groove 67a and radial passage 68in the output shaft. The communication between the chamber 35b and theelongated groove 62, on the other hand, is provided through the radialport 48b in the cylinder block 31, the semi-circumferential groove 64b,elongated groove 65b and radial port 66b in the sleeve 63, and theannular groove 67b in the output shaft 11, as illustrated in FIG. 1.

The sleeve 63 is further formed with elongated grooves 70a and 70bleading from its radial ports 66a and 66b, respectively. These elongatedgrooves 70a and 70b are enclosed in the transmission housing 12 andcommunicate with fluid supply passages 71a and 71b, respectively, whichare formed in the transmission housing as shown. Though not illustrated,these fluid supply passages 71a and 71b lead from a source or sources offluid under pressure through suitable oneway check valves so that theloss in the fluid in the above described hydrostatic circuit resultingfrom the possible leakage of the fluid therefrom is compensated for.These passages 71a and 71b are usually further provided with suitablerelief valves, not shown, which are adapted to prevent the fluidpressure in the hydrostatic circuit from being excessively elevatedduring operation. Such, however, is rather immaterial to theconstruction of the transmission system herein disclosed so that nodetailed discussion will be incorporated.

In order to prevent production of noises otherwise produced in thedirect drive condition of the transmission system and to minimizeleakage of the fluid from the first and third fluid displacement units,it is preferable that valve means be provided for the purpose ofblocking entry of the highly pressurized fluid into the cylindricalchambers of the first and third fluid displacement units, as previouslymentioned.

To accommodate such valve means, the axial bores 54 and 55 in the inputand output shafts have extensions which are formed, by way of example,as stepped larger-diameter bores 72 and 73. These larger-diameter bores72 and 73 are closed at their outer ends by closure plugs 74 and 75which are hermetically fitted to the inner peripheral walls of the inputand output shafts 10 and 11, respectively.

The larger-diameter bore 72 has accommodated therewithin a piston valve76 projecting into the axial bore 54 in the input shaft 10. The pistonvalve 76 has a flange 77 fitting on the inner peripheral wall of theinput shaft 10 so as to divide the larger-diameter bore 72 to aconstant-pressure chamber 78 and a variablepressure chamber 79. Theconstant-pressure chamber 78 is defined by the flange 77 of the pistonvalve 76 and the closure plug 74 and communicates with the axial bore 54in the input shaft 10 through a passage 80 which is formedlongitudinally in the piston valve 76. The variable-pressure chamber 79,on the other hand, communicates with a drain passage 81 through a port82 formed in the inner peripheral wall of the input shaft 10 and anannular groove 83 formed in the outer peripheral wall of the inputshaft. The drain passage 81 is associated with the fluid supply passage71a through a suitable control valve, not shown. This control valve isadapted to open the drain passage 81 for draining the fluid in thevariable-pressure chamber 79 during the direct drive condition and toprovide fluid communication between the passages 71a and 81 forreplenishing the chamber 79 with the fluid from the fluid supply passage710 during the modes of operation excepting the direct drive condition.Suitable spring means and as a compression spring 84 is mounted withinthe variablepressure chamber 79, seated at one end on the flange 77 ofthe piston valve 76 and at the other on a stepped portion, not numbered,of the inner peripheral wall of the input shaft 10, as illustrated. Thecompression spring 84 thus biases the piston valve 76 toward the closureplug 74. The length of the piston valve 76 is such that the piston valveis capable of closing the radial passage 57 in the input shaft 10 whenit is moved away from the closure plug 74 against the action of thecompression spring 84.

Likewise, the larger-diameter bore 55 in the output shaft 11 hasaccommodated therein a piston valve 85 which is movable toward and awayfrom the closure plug 75 and which projects at its leading end into theaxial bore 55 in the output shaft. The piston valve 85 has a flange 86dividing the larger-diameter bore 73 into constant pressure andvariable-pressure chambers 87 and 88, respectively, which arehermetically isolated from each other. The constant-pressure chamber 87communicates with the axial bore 55 in the output shaft 11 through apassage 89 axially formed in the piston valve 85. The variable-pressurechamber 88, on the other hand, communicates with a drain passage 90through a port 91 formed in the inner peripheral wall of the outputshaft 11 and an annular port 92 formed in the outer peripheral wall ofthe output shaft. Similarly to the drain passage 81 from the input shaft10, this drain passage 90 is associated with the fluid supply passage71a through a suitable control valve, not shown, which is adapted toopen the drain passage 90 under the direct drive condition and toprovide fluid communication between the passages 71a and 90 forsupplying the pressurized fluid to the variable-pressure chamber 88. Thevariable-pressure chamber 88 has mounted therein a compression spring 93which acts to bias the piston valve 85 toward the closure plug 75 sothat the radial passage 68 in the output shaft 11 is kept open duringthe modes of operation excepting the direct drive condition.

When, in operation, the cylinder block of any of the fluid displacementunits is driven for rotation, the ball piston elements are radiallymoved within the respective cylindrical chambers toward and away fromthe axis of rotation of the cylinder block provided a certain degree ofeccentricity is established between such axis and the inner cam surfaceof the associated cam ring. If, for instance, the cam ring 21 of thefirst fluid displacement unit 13 is adjusted to be eccentric withrespect to the axis of rotation of the cylinder block 16 of the unit,the ball piston element 20a will be moved inwardly and the ball pistonelement 20b which is diametrically opposed to the former will be movedoutwardly. Such movements of the diametrically opposed ball pistonelements 20a and 20b are reversed if the cam ring 21 is moved in areverse direction so as to establish reversed eccentricity with respectto the axis of rotation of the cylinder block 16. The inward movement ofany of the ball piston elements is accompanied by forced displacement ofthe fluid from the cylindrical chamber for the particular ball pistonelement, thereby giving rise to the fluid pressure in the axial bore 54in the input shaft. The increased fluid pressure is carried to thecylindrical chambers of one or both of the other fluid displacementunits, depending upon the ratio of the displacements which is instantlyestablished between the three fluid displacement units. If, in thisinstance, the cam ring of the variable-displacement unit which may bethe first or third fluid displacement unit is adjusted to be inconcentric alignment with the axis of rotation of the associatedcylinder block, then the delivery of the fluid from the particular unitis interrupted so that the unit remains inoperative on the remainingfluid displacement units. If, moreover, the displacements of any two ofthe fluid displacement units are equalized, the changes in the fluidpressure in the two units are compensated for each other so that thesetwo units are inoperative on the remaining fluid displacement unit.Thus, the variation in the ratios between the displacements of the threedifferent fluid displacement units as caused through adjustment of therelative positions of the cam rings of the first and third fluiddisplacement units establishes various modes of operation in thetransmission system such as the neutral, underdrive, decelerating,direct drive, overdrive and reverse drive conditions.

To assist in the clear understanding of the operation of the hydrostaticpower transmission system shown in FIG. 1, reference is now made to FIG.2 in which the rotary motions of the cylinder blocks of the fluiddisplacement units are translated, on a simulatory basis, intorectilinear movements of linear motion pistons.

The input and output shafts 10 and 11 for the actual power transmissionsystem are herein simulated as linear motion input and output shafts and110, respectively, which are movable back and forth relative to atransmission housing 120. The first, second and third fluid displacementunits 13, 14 and 15 are translated into corresponding hydraulic units130, 140 and 150, respectively. The first hydraulic unit includes alinear motion piston 160 which is representative of the cylinder block16 and a cylinder 200 which is representative of the ball pistonelements 20a, 20b, and the associated cam ring 21. The linear motion 160is connected to the input shaft 100 through a piston rod, not numbered,and moved forwardly in the cylinder 200. The cylinder 200 is internallydivided by the piston 200 into chambers 190a and 190k which arerepresentative of the cylindrical chambers 19a and 19b. The secondhydraulic unit consists of a linear motion piston 260 representative ofthe cylinder block 26 of the second fluid displacement unit 14 and acylinder 280 representative of a combination of the ball piston elements28a, 28b, and the associated cam ring 29. The linear motion piston 260is connected to the piston of the first hydraulic unit 130 and dividesthe cylinder 280 into chambers 270a and 270b representing thecylindrical chambers 27a and 27b, respectively, in the cylinder block26. Likewise, the third hydraulic unit 150 is made up of a linear motionpiston 310 corresponding to the cylinder block 31 of the thirddisplacement unit 15 and a cylinder 360 standing for a combination ofthe ball piston elements 36a, 36b, and the cam ring 37. The linearmotion piston 310 is connected to the cylinder 200 of the secondhydraulic unit 140 through a piston rod, not numbered, and divided thecylinder 360 into chambers 350a and 35% which correspond to thecylindrical chambers 35a and 35b, respectively, in the cylinder block31. The piston 310 is further connected to the output shaft 1 10. Thecylinders 200 and 360 of the first and third hydraulic units 140 and150, respectively, are held stationary relative to the transmissionhousing 120, while the cylinder 280 of the second hydraulic unit 140 isaxially movable relative to the transmission housing.

The chamber a of the first hydraulic unit 130 communicates with thechamber 270a of the second hydraulic unit 140 through a port 460a, apassage 540 and a port 470a. The port 460a thus represents a fluid lineconsisting of the radial port 46a in the cylinder block 16, thesemi-circumferential groove 50a, elongated groove 51a and radial port52a in the sleeve 49, and the annular groove 56a and radial passage 57in the input shaft 10. The passage 540 corresponds to the axial bore 54in the input shaft and a portion of the axial port 55 in the outputshaft 11. The port 470a is representative of a fluid line constituted bythe radial port 47a in the cylinder block 26, and thesemi-circumferential groove 59a and radial passage 60 in the outputshaft 11 associated with the second fluid displacement unit 14.

The passage 540 also communicates with the chamber 350a of the thirdhydraulic unit 150 through a passage 550 and a port 480a. The passage550 corresponds to the remaining portion of the axial bore 55 in theoutput shaft. The port 480a represents a fluid line established by theradial port 48a in the cylinder block 31, the semi-circumferentialgroove 64a, elongated groove 65a and radial port 66a in the output shaft11.

The chamber 190b of the first hydraulic unit 130 communicates with thechamber 270b of the second hydraulic unit 140 through a port 460b, apassage 580 and a port 47%. The port 460b is representative of a fluidline established by the radial port 46b in the cylin der block 16, thesemi-circumferential groove 50b, elongated groove 51b and radial port52b in the sleeve 49, and the annular groove 56b in the input shaft 10.The passage 580 corresponds to the elongated groove 58 in the inputshaft and the annular 61 formed between the input and output shaft 10and 11, respectively. The port 47% is representative of a fluid lineconstituted by the radial port 47b in the cylinder block 26, and thesemi-circumferential groove 5% in the output shaft 11 associated withthe second fluid displacement unit 14.

The passage 580 also communicates with the chamber 35Gb of the thirdhydraulic unit 150 through a passage 620 and a port 480b. The passage620 corresponds to the elongated groove 62 in the output shaft 11 whilethe port 48% is representative of a fluid line which is established bythe radial port 48b in the cylinder block 31, the semi-circumferentialgroove 64b, elongated groove 65b and radial port 66b in the sleeve 63and the annular groove 67b in the output shaft 11 associated with thethird fluid displacement unit 15,

The diameters of the cylinders 200, 280 and 360 of the simulatedhydraulic units are assumed to correspond to the displacements of thecorresponding fluid displacement units of the actual transmissionsystem. Thus, diameters of the cylinders 200 and 360 of the first andthird hydraulic units 130 and 150, respectively, are variable while thediameter of the cylinder 280 of the second hydraulic unit 140 is keptconstant.

It is now assumed that the rotational speeds of the input and outputshafts are N, and N respectively, and that the displacements of thefirst, second and third fluid displacement units are V,, V, and Vrespectively, per turn of the cylinder blocks. Here, the values Themodes of operation of the simulated power transmission shown in FIG. 2will be accounted for by this Equation 2 as follows:

NEUTRAL CONDITION If the value V is held at a maximum and concurrentlyV, V,,'then the speed reduction ratio R is an infinity so that theoutput shaft is held at a standstill. If, thus, the diameter of thecylinder 200 of the first hydraulic unit 130 is adjusted to be equal tothe fixed diameter of the cylinder 280 of the second hydraulic unit, thefluid discharged from the chamber 270a as a result of the movement ofthe input shaft is totally passed to the chamber 190a through thepassage 540 with the result that no driving power is imparted to theoutput shaft 110. Since, in this condition, the third hydraulic unit 150fails to deliver or receive the fluid pressure, a braking action with arelatively great magnitude is applied to the output shaft 110.

UNDERDRIVE CONDITION If the value V is invariably held at a maximum andat the same time a relation 0 V, V holds, then the speed reduction ratioR diminishes and the value N increases as the value V, approaches zero.This means that, if the diameter of the cylinder 200 of the firsthydraulic unit 130 is smaller than the diameter of the cylinder 280 ofthe second hydraulic unit 140, the chamber 190a can not afford toreceive the total amount of the fluid discharged from the chamber 270aso that the fluid from the chamber 270a is partly circulated to thechamber 350a of the third hydraulic unit 150 through the passage 550.The output shaft is consequently moved rightwardly of the drawing.

LOW-SPEED CONDITION When V, 0, the speed reduction ratio R is expressedas:

If, therefore, the value V, is decreased with V, 0, then the speedreduction ratio R decreases and the value N increases as the value Mincreases. In other words, if the inside diameter of the cylinder 200 ofthe first hydraulic unit is in agreement with the piston rod for thepiston 160 so that the chamber 190a has a zero capacity, then the fluiddischarged from the 270a of the second hydraulic unit is totally passedto the chamber 350a of the third hydraulic unit through the passage 550.Thus, the piston 310 of the third bydraulic unit 150 is subjected notonly to a reaction of the cylinder 280 to the fluid pressure in thechamber 270a but to the fluid pressure thus built up in the chamber350a.

The output shaft 110 is, therefore, decelerated at a speed considerablyhigher than during the underdrive condition if the displacement of thefirst hydraulic unit 130 is interrupted and at the same time thedisplacement of the third hydraulic unit 150 is increased.

. DIRECT DRIVE CONDITION If V, V, 0, then R 1 so that the output shaftis driven at the same speed as the input shaft. If, thus, the insidediameters of the cylinders 200 and 360 of the first and third hydraulicunits 130 and 150 are equal to the diameters of the piston rods for thereciprocating positors 160 and 310, respectively the fluid in the 270ais constrained therewithin with the result that a hydraulic lockcondition is built up. The input shaft 100, second hydraulic unit 140and output shaft 110 consequently act as an integral unit so that theoutput shaft 110 is driven at the same speed as the input shaft 100.

OVERDRIVE CONDITION If V and V, 0, then the second term of Equation 2assumes a minus value which is greater than 1. Accordingly, the speedreduction ratio R is smaller than 1 so that the value N is greater thanthe value N To make the value V minus is to have the passages 540 and580 reversed from each other. The passage 540 leading from the chamber190a of the first hydraulic unit 130 is open to the chamber 27% of thesecond hydraulic unit 140, while the passage 580 is open to the chamber270a. The diameter of the cylinder 360 is equal to the diameter of thepiston rod for the reciprocating piston 310 so that the displacement ofthe third hydraulic unit 150 is interrupted. The fluid discharged fromthe chamber 19Gb resulting from the movenent of the input shaft 100 isthus circulated to the chamber 270a of the second hydraulic unit 140.The result is that the output shaft 110 is driven by the power resultingfrom the movement of the piston 260 in the cylinder 280 plus the fluidpower resulting from the delivery of the fluid from the chamber 19% tothe chamber 270a. The output shaft 110 is accordingly driven at a higherspeed than the input shaft 100, whereby the overdrive condition isestablished.

REVERSE DRIVE CONDITION If the value V is held at a maximum andconcurrently V, V then the second term of Equation 2 assumes a minusvalue which in this instance is smaller than --I. The speed reductionratio R is consequently minus so that the values N and N have oppositesigns, viz., the output shaft is driven in an opposite direction to theinput shaft. In other words, if the diameter of the cylinder 200 of thefirst hydraulic unit 130 is larger than the diameter of the cylinder 280of the second hydraulic unit 140 and if the diameter of the cylinder 360of the third hydraulic unit 150 is increased to a maximum, then thefluid discharged from the chamber 19% of the first hydraulic unit 130 asa result of the movement of the input shaft 100 is passed partly to thechamber 27% of the second hydraulic unit 140 through the passage 580 andpartly to the chamber 35Gb of the third hydraulic unit 150 through thepassage 620 branched from the passage 580. The piston 310 of the thirdhydraulic unit 150 is accordingly moved leftwardly of the drawing sothat the output shaft 110 is driven in the reverse direction.

It will now be understood from the above discussion that the secondhydraulic unit 140 operates in a manner that action and reaction areexercised between the piston 260 and the cylinder 280 and that the unitis movable in its entirety. The second hydraulic unit is thus capable oftransmitting the power in a mechanical fashion.

Under the underdrive condition in which the displacement of the firsthydraulic unit is smaller than the displacement of the second hydraulicunit, the fluid displaced from the second displacement unit is dividedinto two portions. One of such portions of the fluid is passed to thefirst hydraulic unit so as to add to the driving power for the pistontherein and the other of the portions is passed to the third hydraulicunit so as to add to the driving power for the piston therein. Thetransmission therefore operates in a manner to have its fluidsplit-power into input and output components under the underdrivecondition.

Under the condition at a relatively high speed, the third hydraulic unitreceives not only the fluid power resulting from the delivery of thefluid from the second hydraulic unit but the power mechanicallytransmitted thereto by the reaction of the cylinder of the secondbydraulic unit to the movement of the piston thereof. The power suppliedis thus split into two output components.

When the overdrive condition is established in which the fluid in thefirst hydraulic unit is circulated to the second hydraulic unit, thepiston of the third hydraulic unit is subjected to the mechanical powerresulting from the reaction of the cylinder of the second hydraulic unitto the piston thereof which is directly moved by the input shaft and tothe fluid power resulting from the fluid passed to the second hydraulicunit. The power supplied to the transmission is in this manner splitinto two output components.

In the reverse drive condition in which the displacement of the firsthydraulic unit is smaller than that of the second hydraulic unit, thefluid power is split into two input components, one being passed to thesecond hydraulic unit and the other to the third hydraulic unit.

The operation of the piston valves 76 and 85 will now be described withreference to FIG. I. As formerly mentioned, these piston valves 76 and85 are intended to block entry of the pressurized fluid into the firstand third fluid displacement units during the direct drive condition inwhich the displacements of these fluid displacement units areinterrupted to cause the input and output shafts 10 and 11 and thetransmission system to operate as an integral unit.

When, thus, the first and third fluid displacement units 13 and 15become inoperative to deliver the fluid, then the previously mentionedcontrol valves associated with the drain passages 81 and 90 are actuatedto open these drain passages. The fluid staying in the variable-pressurechambers 79 and 88 of the largerdiameter bores 72 and 73, respectively,is in this manner drained off. The piston valves 76 and 85 areconsequently moved against the actions of the compression springs 84 and93 away from the closure plugs 74 and 75 by the pressure obtaining inthe constant-pressure chambers 78 and 87, respectively. The piston-valve'76 and protrudes deeper into the axial bores 54 and 55 in the input andoutput shafts 10 and 11, respectively, thereby closing the radialpassages 57 and 68 leading from the axial bores 54 and 55, respectively.The pressurized fluid in the axial bores 54 and 55 is consequentlyprevented from entering the first and third fluid displacement units 13and 15. Thus, even though the radial ports in the cylinder blocks 16 and31 are periodically opened due to the variation in the angular positionsof the cylinder blocks relative to the sleeves 49 and 63 as thesecylinder blocks rotate about their axes, the pulsating'fluid pressure inthe axial bores 54 and 55 is not carried to the cylindrical chambers inthe cylinder blocks. This is useful in avoiding production of the noisesin the first and third fluid displacement units and in minimizing theleakage of the fluid from the cylinder blocks of the units.

When the direct drive condition is to be terminated, the control valvesfor the drain passages 81 and 90 are shifted to positions to providecommunication between these passages and the fluid supply passage 71a.The variable-pressure chambers 79 and 88 are consequently replenishedwith the fluid under pressure so that the piston valves 76 and 85restores their initial positions with the fluid in the constant-pressurechambers 78 and 87 recirculated to the axial bores 54 and 55 through thepassages 80 and 89, respectively.

The hydrostatic power transmission system which has thus far beendescribed in detail has outstanding features which are enumeratedasfollows:

a. Simplicity of construction.

b. Capability of dealing with powers of relatively great magnitudes inspite of the small-sized construction.

0. Ease of operation and maintenance servicing.

d. Speeds and torques varied over a stepless range in either directionfrom zero to maximum.

e. Split-power characteristics in which the supplied power istransmitted partly in a mechanical fashion and partly in a hydraulicfashion in various modes of operation, thereby significantly adding tothe transmission efficiency. I

f. Void of mechanical clutches and/or geared reduction mechanisms.

g. Minimized loss in the transmission efficiency during the direct drivecondition in which only the second fluid displacement is operative toestablish the hydraulic lock condition.

h. Minimized noises and leakage of the fluid from the first and thirdfluid displacement units during the direct drive condition in which thedisplacement of these units are interrupted.

A second preferred embodiment of the hydrostatic power transmission inaccordance with this invention is now illustrated in FIG. 3. Thetransmission system shown in FIG. 3 is similar to the system shown inFIG. 1 in that it uses first and second fluid displacement units asvariable-displacement and constantdisplacement units, respectively.Different from the system of FIG. 1, the modified system includes athird fluid displacement unit acting as a constant displacement unit.The third fluid displacement unit, which as such is designated by newreference numeral 94, thus includes a cam ring 95 which is stationaryrelative to the transmission housing 12. This cam ring 95 may be eithersecurely connected to the transmission housing 12 or formed as integralwith the transmission housing. Similarly to the cam ring 37 of the thirdfluid displacement unit of the first embodiment, the cam ring 95 has acircular inner cam surface with the ball piston elements 36a, 36b,respectively received within the cylindrical chambers 35a,'35b, in thecylinder block 31 are in frictional engagement. Designated by 96 is anannular groove formed in the cam surface of the cam ring 95.

Furthermore, the transmission system shown in FIG. 3 is void of thepiston valves 76 and 85 which are used in the system shown in FIG. 1.Thus, the axial bores 54 and 55 in the input and output shafts 10 and 11are closed at their opposed ends which are close to the ends of theflrst and third fluid displacement units, respectively, as shown.

Thus, in the transmission system shown in FIG. 3, the speed reductionratio can be varied by varying the displacement of the first fluiddisplacement unit 13 through adjustment of the angular position of itscam ring 21 with respect to the associated cylinder block 16. The speedreduction ratio R and the rotational speed of the output shaft 11 areaccordingly expressed R 2 3) 2 1) and q- 2 2 r V2 a n qwhere the valuesV and V are constants. These equations are graphically represented bycurves R and a line N in FIG. 4, in which the rotational speed N of theinput shaft is assumed to be a constant. The axis a'a of abscissa standsfor the variation in the displacement V of the first fluid displacementunit while the axis b'-b of ordinate stands for the variations in thespeed reduction ratio R and the rotational speed N of the output shaft.The speed reduction ratio R is thus represented by a hyperbola havingasymptotes a'a and cc which perpendicularly intersects the axis a'a ofabscissa at point d which is indicative of the value V The speedreduction ratio R thus assumes a plus value if the value V is smallerthan the value V and a minus value if the value V is greater than thevalue V The rotational speed N of the output shaft is represented by astraight line passing through the point d at which the value V equalsthe value V As a consequence, the output shaft is brought to astandstill if the displacement V of the first fluid displacement unit isadjusted to the displacement V of the second fluid displacement unit,whereby the neutral condition is established. If the displacement V, isincreased beyond the displacement V then the output shaft is driven in adirection reverse to the direction of rotation of the input shaft,thereby establishing the reverse drive condition. If, conversely, thedisplacement V becomes smaller than the displacement V then the outputshaft is driven in the same direction as the input shaft so as toestablish the underdrive or lowspeed condition. If, furthermore, thedirection of the displacement of the first fluid displacement unit isreversed and if the absolute values of the displacements of the firstand third fluid displacement units are equalized to each other so that V--V;,, then the output shaft is driven at the same speed and in the samedirection as the input shaft. This establishes the direct drivecondition.

If the displacement in either direction, of the first fluid displacementunit is greater than the displacement of the third fluid displacementunit so that V, V then the output shaft outspeeds the input shaft toestablish the overdrive condition.

The features of the hydrostatic power transmission thus constructed andoperating are summarized as follows:

a. Simplicity of constuction.

b. Split-power characteristics in which the supplied driving power issplit into mechanical and fluid powers or into input or output componentor a combination of input and output components.

c. The power input is directly transmitted in part to the first fluiddisplacement unit and in part to the second fluid displacement unit. Thesecond fluid displacement unit drives the output shaft by its reactionresulting from the fluid power transmitted thereto. No loss in themechanical power is consequently invited in the power transmission bythe second fluid displacement unit. On the other hand, the combinedfluid pressures built up by the first and second fluid displacementunits act on the third fluid displacement unit so as to hydrostaticallyproduce a torque therein. A loss in the hy draulic power is thus invitedin the third fluid displacement unit. This loss, however, results indrops of the transmission efficiency which are comparable to thoseexperienced in hydrostatic power transmissions in which the power inputis transmitted totally in a hydrostatic fashion. Such drops in thetransmission efficiency can be offset by the increased transmissionefficiency attained by the mechanical power transmission by the secondfluid displacement unit. As a consequence, the total transmissionefficiency which is attained in the transmission system as a whole ismaintained at a considerably elevated level.

d. Simplicity of operation, especially in regulating the displacementratios between the three fluid displacement units. Such ratios can bevaried simply by adjusting the relative position of only one adjustablecam ring of the first fluid displacement unit.

What is claimed is:

1. A hydrostatic power transmission system for delivering a drivingtorque from an input shaft to an output shaft which is in line with theinput shaft comprising, in combination, first, second and third fluiddisplacement units each comprising a rotatable cylinder block coaxialwith the input and output shafts and a cam ring positioned around saidcylinder block, the cylinder blocks of the first and second fluiddisplacement units being rotatable with said input shaft, the cam ringof the second fluid displacement unit and the cylinder block of thethird fluid displacement unit being in driving engagement with saidoutput shaft, each of said cylinder blocks having a plurality ofcylinders which are directed toward an axis of rotation of the cylinderblock and a plurality of piston elements which are respectively receivedwithin said cylinders and movable toward and away from said axis, saidpiston elements being in frictional engagement with said cam surface ofthe associated cam ring, the cam ring of the first fluid displacementunit being movable over the associated cylinder block in a planesubstantially transverse to the axis of rotation of the cylinder blockfor providing controlled degrees of eccentricity between the cam surfaceof the particular cam ring and the axis-of rotation of the associatedcylinder block, actuating means for moving the cam ring of said firstfluid displacement unit in said plane for varying the displacement ofthe first fluid displacement unit, the cam ring of the second fluiddisplacement unit being held stationary for keeping constant thedisplacement of the second fluid displacement unit, and fluid pressuredistribution passage means providing controlled fluid communicationbetween the cylinders of said first, second and third fluid displacementunits for radially outwardly biasing said piston elements within each ofthe cylinders, said first fluid displacement unit being operable todisplace the fluid in either direction from zero to maximum.

2. A hydrostatic power transmission system according to claim 1, inwhich the maximum displacement per turn of the cylinder block of thefirst displacement unit being smaller than the displacement per turn ofthe cylinder block of said second fluid displacement unit, said thirdfluid displacement unit being operable to displace the fluid in a fixeddirection.

3. A hydrostatic power transmission system according to claim 1, inwhich the cam ring of said first fluid displacement unit is movedrelative to the associated cylinder block to a position in which thedisplacement of the first fluid displacement unit is smaller than thedisplacement of the second fluid displacement unit for causing saidoutput shaft to be driven at a speed relatively lower than and in thesame direction as said input shaft to provide an underdrive condition.

4. A hydrostatic power transmission system according to claim 1, inwhich the cam ring of the first fluid displacement unit is movedrelative to the associated cylinder block to a position in which thedisplacement thereof equals the displacement of the second fluiddisplacement unit for causing said output shaft to stop to provide aneutral condition.

5. A hydrostatic power transmission system according to claim 1, inwhich the cam ring of said third fluid displacement unit is movable overthe associated cylinder block in a plane substantially transverse to anaxis of rotation of the cylinder block for providing controlled degreesof eccentricity between the cam surface of the cam ring and the axis ofrotation of the cylinder block, said system further comprising actuatingmeans for displacing the cam ring of said third fluid displacement unitin said plane for varying the displacement of the third fluiddisplacement unit, said cam ring of the third fluid displacement unitbeing moved relative to the associated cylinder block to a positionproviding a maximum displacement of the third fluid displacement unitduring neutral, underdrive and reverse drive conditions.

6. A hydrostatic power transmission system according to claim 5, inwhich the cam ring of said first fluid displacement unit is adjusted tobe in concentric alignment with the associated cylinder block tointerrupt the displacement of the first fluid displacement unit to causesaid output shaft to be driven at a speed lower than said input speedbut higher than the speed of rotation of the output speed during theunderdrive condition, wherein the cam ring of said third fluiddisplacement unit is moved relative to the associated cylinder block toa position in which the displacement of the third fluid displacementunit is decreased for establishing a low-speed condition.

7. A hydrostatic power transmission system according to claim 5, inwhich the cam rings of the first and third fluid displacement units aremoved relative to the associated cylinder blocks to be in concentricalignment with these cylinder blocks to interrupt the displacements ofthe two units concurrently for causing said output shaft to be driven atthe same speed and in the same direction as said input shaft toestablish a direct drive condition.

8. A hydrostatic power transmission system according to claim 5, inwhich the cam ring of the first fluid displacement unit is movedrelative to the associated cylinder block to a position in which thedirection of the displacement thereof is reversed and concurrently thecam ring of the third fluid displacement unit is moved relative to theassociated cylinder block to be in concentric alignment with thecylinder block to cause said output shaft to be driven at a higher speedthan said input shaft to establish an overdrive condition.

9. A hydrostatic power transmission system according to claim 5, furthercomprising valve means operable to block entry of pressurized fluid intosaid first and third fluid displacement units during a direct drivecondition.

10. A hydrostatic power transmission system according to claim 1, inwhich the cam ring of the third fluid displacement unit is heldstationary.

11. A hydrostatic power transmission system according to claim 10, inwhich the cam ring of the first fluid displacement unit is movedrelative to the associated cylinder block to a position in which thedisplacement of the first fluid displacement unit is smaller than thedisplacement of the second fluid displacement unit for causing saidoutput shaft to be driven in the same direction as said input shaft.

12. A hydrostatic power transmission system according to claim 10, inwhich the cam ring of said first fluid displacement unit is movedrelative to the associated cylinder block to a position in which thedisplacement of the first fluid displacement unit is larger than thedisplacement of the second fluid displacement unit for causing saidoutput shaft to be driven in a reverse direction to said input shaft.

13. A hydrostatic power transmission system according to claim 10, inwhich the cam ring of said first fluid displacement unit is movedrelative to the associated cylinder block to a position in which thedisplacement in a reversed direction of the first fluid displacementunit equals the displacement of the third fluid displacement unit.

14. A hydrostatic power transmission system according to claim 10, inwhich the cam ring of the first fluid displacement unit is movedrelative to the associated cylinder block to a position in which thedisplacement in either direction of the first fluid displacement unit isgreater than the displacement of the third fluid displacement unit.

15. A hydrostatic power transmission system according to claim 1,further comprising a first stationary sleeve which is interposed betweensaid input shaft and the cylinder block of said first fluid displacementunit and a second stationary sleeve which is interposed between saidoutput shaft and the cylinder block of said third fluid displacementunit, said fluid pressure distribution passage being formed by radialports formed in inner peripheral walls of the cylinder blocks of saidfirst, second and third fluid displacement units and communicating withsaid cylinders therein, a pair of diametrically spacedsemi-circumferential grooves formed in an outer peripheral wall of saidstationary sleeve and aligned with said radial ports in the cylinders ofsaid first fluid displacement unit, apair of elongated grooves formed inthe outer peripheral wall of said first stationary sleeve andcommunicating with said semi-circumferential grooves respectively, apair of radial ports formed in an inner peripheral wall of said firststationary sleeve communicating with said elongated groovesrespectively, a pair of spaced annular grooves formed in an outerperipheral wall of said input shaft and constantly communicating withsaid radial ports in said first stationary sleeve, an axial bore formedin said input shaft and opened at its terminal end, a radial passageformed in said input shaft providing communication between one of saidannular grooves, an elongated groove formed in the outer peripheral wallof said input shaft and terminating at said terminal end of the inputshaft, said elongated groove in said inner shaft communicating with theother of said annular grooves, an annular groove formed betweenconjoined ends of the input and output shafts, an axial bore formed insaid output shaft and opened at its starting end, the axial bores insaid input and output shafts being in communication with each other andsubstantially coextensive with said first, second and third fluiddisplacement units, a first radial passage formed in said output shaftand communicating with said axial bore in the output shaft, a pair ofdiametrically spaced semi-circumferential ports formed in an outerperipheral wall of said output shaft and aligned with said radial portsin the cylinders of said second fluid displacement unit, one of whichsemi-circumferential ports communicates with said first radial passageand the other of which semi-circumferential ports communicates with saidannular groove between said joined ends of the input and output shafts,an elongated groove formed in the outer peripheral wall of said outputshaft and communicating with said other of said semi-circumferentialports in the input shaft, a second radial passage formed in said outputshaft and communicating with said axial bore in the output shaft, saidradial bore in said input shaft and the first and second radial passagesin said output shaft being aligned with each other and diametricallyopposed to the elongated grooves in the outer peripheral walls of theinput and output shafts, a pair of spaced annular grooves formed in theouter peripheral wall of said output shaft and communicatingrespectively with said second radial passage and said elongated groovein said output shaft, a pair of radial ports formed in an innerperipheral wall of said second stationary sleeve and constantlycommunicating respectively with said annular grooves in said outputshaft, a pair of elongated grooves formed in an outer peripheral wall ofsaid second stationary sleeve and communicating respectively with saidradial ports therein, a pair of diametrically spacedsemi-circumferential ports formed in the outer peripheral wall of saidsecond stationary sleeve and communicating respectively with saidelongated grooves therein, said semi-circumferential ports in the secondstationary sleeve being aligned with said radial ports in the cylindersof said third fluid displacement unit, and a pair of valved fluid supplypassages communicating with said radial ports in said second stationarysleeve and with a source of pressurized fluid.

16. A hydrostatic power transmission system according to claim 15,further comprising valve means for closing said radial passage in saidinput shaft and said second radial passage in said output shaft when thecam rings of the first and third fluid displacement units are movedrelative to the associated cylinder blocks to be in concentric alignmentwith the cylinder blocks for interrupting the displacements of the twofluid displacement units to establish a direct drive condition.

17. A hydrostatic power transmission system according to claim 16, inwhich said valve means comprises a pair of piston valves axially movablyaccommodated within first and second larger-diameter bores extendingfrom opposed ends of said axial bores in said input and output shaftsrespectively, said larger-diameter bores being closed at their outerends by closure plugs respectively, said valve pistons projecting attheir leading ends into said axial bores in said input and output shaftsrespectively, each of said valve pistons having a flange by which thelarger-diameter bore receiving the piston therewithin is divided intoconstant-pressure and variable-pressure chambers and a passage formedtherein for providing communication between the constantand drainpassages for causing said valve pistons be movedtoward said closureplugs by a fluid pressure passed to the variable-pressure chambers andfor opening said drain passages during said direct drive condition tocause said valve pistons to move to positions to close said radialpassage in said input shaft and said first radial passage in said outputshaft.

1. A hydrostatic power transmission system for delivering a drivingtorque from an input shaft to an output shaft which is in line with theinput shaft comprising, in combination, first, second and third fluiddisplacement units each comprising a rotatable cylinder bLock coaxialwith the input and output shafts and a cam ring positioned around saidcylinder block, the cylinder blocks of the first and second fluiddisplacement units being rotatable with said input shaft, the cam ringof the second fluid displacement unit and the cylinder block of thethird fluid displacement unit being in driving engagement with saidoutput shaft, each of said cylinder blocks having a plurality ofcylinders which are directed toward an axis of rotation of the cylinderblock and a plurality of piston elements which are respectively receivedwithin said cylinders and movable toward and away from said axis, saidpiston elements being in frictional engagement with said cam surface ofthe associated cam ring, the cam ring of the first fluid displacementunit being movable over the associated cylinder block in a planesubstantially transverse to the axis of rotation of the cylinder blockfor providing controlled degrees of eccentricity between the cam surfaceof the particular cam ring and the axis of rotation of the associatedcylinder block, actuating means for moving the cam ring of said firstfluid displacement unit in said plane for varying the displacement ofthe first fluid displacement unit, the cam ring of the second fluiddisplacement unit being held stationary for keeping constant thedisplacement of the second fluid displacement unit, and fluid pressuredistribution passage means providing controlled fluid communicationbetween the cylinders of said first, second and third fluid displacementunits for radially outwardly biasing said piston elements within each ofthe cylinders, said first fluid displacement unit being operable todisplace the fluid in either direction from zero to maximum.
 2. Ahydrostatic power transmission system according to claim 1, in which themaximum displacement per turn of the cylinder block of the firstdisplacement unit being smaller than the displacement per turn of thecylinder block of said second fluid displacement unit, said third fluiddisplacement unit being operable to displace the fluid in a fixeddirection.
 3. A hydrostatic power transmission system according to claim1, in which the cam ring of said first fluid displacement unit is movedrelative to the associated cylinder block to a position in which thedisplacement of the first fluid displacement unit is smaller than thedisplacement of the second fluid displacement unit for causing saidoutput shaft to be driven at a speed relatively lower than and in thesame direction as said input shaft to provide an underdrive condition.4. A hydrostatic power transmission system according to claim 1, inwhich the cam ring of the first fluid displacement unit is movedrelative to the associated cylinder block to a position in which thedisplacement thereof equals the displacement of the second fluiddisplacement unit for causing said output shaft to stop to provide aneutral condition.
 5. A hydrostatic power transmission system accordingto claim 1, in which the cam ring of said third fluid displacement unitis movable over the associated cylinder block in a plane substantiallytransverse to an axis of rotation of the cylinder block for providingcontrolled degrees of eccentricity between the cam surface of the camring and the axis of rotation of the cylinder block, said system furthercomprising actuating means for displacing the cam ring of said thirdfluid displacement unit in said plane for varying the displacement ofthe third fluid displacement unit, said cam ring of the third fluiddisplacement unit being moved relative to the associated cylinder blockto a position providing a maximum displacement of the third fluiddisplacement unit during neutral, underdrive and reverse driveconditions.
 6. A hydrostatic power transmission system according toclaim 5, in which the cam ring of said first fluid displacement unit isadjusted to be in concentric alignment with the associated cylinderblock to interrupt the displacement of the first fluid diSplacement unitto cause said output shaft to be driven at a speed lower than said inputspeed but higher than the speed of rotation of the output speed duringthe underdrive condition, wherein the cam ring of said third fluiddisplacement unit is moved relative to the associated cylinder block toa position in which the displacement of the third fluid displacementunit is decreased for establishing a low-speed condition.
 7. Ahydrostatic power transmission system according to claim 5, in which thecam rings of the first and third fluid displacement units are movedrelative to the associated cylinder blocks to be in concentric alignmentwith these cylinder blocks to interrupt the displacements of the twounits concurrently for causing said output shaft to be driven at thesame speed and in the same direction as said input shaft to establish adirect drive condition.
 8. A hydrostatic power transmission systemaccording to claim 5, in which the cam ring of the first fluiddisplacement unit is moved relative to the associated cylinder block toa position in which the direction of the displacement thereof isreversed and concurrently the cam ring of the third fluid displacementunit is moved relative to the associated cylinder block to be inconcentric alignment with the cylinder block to cause said output shaftto be driven at a higher speed than said input shaft to establish anoverdrive condition.
 9. A hydrostatic power transmission systemaccording to claim 5, further comprising valve means operable to blockentry of pressurized fluid into said first and third fluid displacementunits during a direct drive condition.
 10. A hydrostatic powertransmission system according to claim 1, in which the cam ring of thethird fluid displacement unit is held stationary.
 11. A hydrostaticpower transmission system according to claim 10, in which the cam ringof the first fluid displacement unit is moved relative to the associatedcylinder block to a position in which the displacement of the firstfluid displacement unit is smaller than the displacement of the secondfluid displacement unit for causing said output shaft to be driven inthe same direction as said input shaft.
 12. A hydrostatic powertransmission system according to claim 10, in which the cam ring of saidfirst fluid displacement unit is moved relative to the associatedcylinder block to a position in which the displacement of the firstfluid displacement unit is larger than the displacement of the secondfluid displacement unit for causing said output shaft to be driven in areverse direction to said input shaft.
 13. A hydrostatic powertransmission system according to claim 10, in which the cam ring of saidfirst fluid displacement unit is moved relative to the associatedcylinder block to a position in which the displacement in a reverseddirection of the first fluid displacement unit equals the displacementof the third fluid displacement unit.
 14. A hydrostatic powertransmission system according to claim 10, in which the cam ring of thefirst fluid displacement unit is moved relative to the associatedcylinder block to a position in which the displacement in eitherdirection of the first fluid displacement unit is greater than thedisplacement of the third fluid displacement unit.
 15. A hydrostaticpower transmission system according to claim 1, further comprising afirst stationary sleeve which is interposed between said input shaft andthe cylinder block of said first fluid displacement unit and a secondstationary sleeve which is interposed between said output shaft and thecylinder block of said third fluid displacement unit, said fluidpressure distribution passage being formed by radial ports formed ininner peripheral walls of the cylinder blocks of said first, second andthird fluid displacement units and communicating with said cylinderstherein, a pair of diametrically spaced semi-circumferential groovesformed in an outer peripheral wall of said stationary sleeve and alignEdwith said radial ports in the cylinders of said first fluid displacementunit, a pair of elongated grooves formed in the outer peripheral wall ofsaid first stationary sleeve and communicating with saidsemi-circumferential grooves respectively, a pair of radial ports formedin an inner peripheral wall of said first stationary sleevecommunicating with said elongated grooves respectively, a pair of spacedannular grooves formed in an outer peripheral wall of said input shaftand constantly communicating with said radial ports in said firststationary sleeve, an axial bore formed in said input shaft and openedat its terminal end, a radial passage formed in said input shaftproviding communication between one of said annular grooves, anelongated groove formed in the outer peripheral wall of said input shaftand terminating at said terminal end of the input shaft, said elongatedgroove in said inner shaft communicating with the other of said annulargrooves, an annular groove formed between conjoined ends of the inputand output shafts, an axial bore formed in said output shaft and openedat its starting end, the axial bores in said input and output shaftsbeing in communication with each other and substantially coextensivewith said first, second and third fluid displacement units, a firstradial passage formed in said output shaft and communicating with saidaxial bore in the output shaft, a pair of diametrically spacedsemi-circumferential ports formed in an outer peripheral wall of saidoutput shaft and aligned with said radial ports in the cylinders of saidsecond fluid displacement unit, one of which semi-circumferential portscommunicates with said first radial passage and the other of whichsemi-circumferential ports communicates with said annular groove betweensaid joined ends of the input and output shafts, an elongated grooveformed in the outer peripheral wall of said output shaft andcommunicating with said other of said semi-circumferential ports in theinput shaft, a second radial passage formed in said output shaft andcommunicating with said axial bore in the output shaft, said radial borein said input shaft and the first and second radial passages in saidoutput shaft being aligned with each other and diametrically opposed tothe elongated grooves in the outer peripheral walls of the input andoutput shafts, a pair of spaced annular grooves formed in the outerperipheral wall of said output shaft and communicating respectively withsaid second radial passage and said elongated groove in said outputshaft, a pair of radial ports formed in an inner peripheral wall of saidsecond stationary sleeve and constantly communicating respectively withsaid annular grooves in said output shaft, a pair of elongated groovesformed in an outer peripheral wall of said second stationary sleeve andcommunicating respectively with said radial ports therein, a pair ofdiametrically spaced semi-circumferential ports formed in the outerperipheral wall of said second stationary sleeve and communicatingrespectively with said elongated grooves therein, saidsemi-circumferential ports in the second stationary sleeve being alignedwith said radial ports in the cylinders of said third fluid displacementunit, and a pair of valved fluid supply passages communicating with saidradial ports in said second stationary sleeve and with a source ofpressurized fluid.
 16. A hydrostatic power transmission system accordingto claim 15, further comprising valve means for closing said radialpassage in said input shaft and said second radial passage in saidoutput shaft when the cam rings of the first and third fluiddisplacement units are moved relative to the associated cylinder blocksto be in concentric alignment with the cylinder blocks for interruptingthe displacements of the two fluid displacement units to establish adirect drive condition.
 17. A hydrostatic power transmission systemaccording to claim 16, in which said valve means comprises a pair ofpiston valves axially movAbly accommodated within first and secondlarger-diameter bores extending from opposed ends of said axial bores insaid input and output shafts respectively, said larger-diameter boresbeing closed at their outer ends by closure plugs respectively, saidvalve pistons projecting at their leading ends into said axial bores insaid input and output shafts respectively, each of said valve pistonshaving a flange by which the larger-diameter bore receiving the pistontherewithin is divided into constant-pressure and variable-pressurechambers and a passage formed therein for providing communicationbetween the constant-pressure chambers of the two larger-diameter boresand said axial bores in said input and output shafts, and a pair ofspring means each mounted within said variable-pressure chambers andbiasing each of said pistons toward each of said closure plugs, thevariable-pressure chambers being respectively in communication with saidfluid supply passages through respective drain passages which areprovided with control valves for usually providing communication betweensaid fluid supply and drain passages for causing said valve pistons bemoved toward said closure plugs by a fluid pressure passed to thevariable-pressure chambers and for opening said drain passages duringsaid direct drive condition to cause said valve pistons to move topositions to close said radial passage in said input shaft and saidfirst radial passage in said output shaft.